Reciprocating high-pressure pumps used in oil and gas fields, particularly those intended for fracking, are usually designed in two sections. The (proximal) power section (herein “power end”) and the (distal) fluid section (herein “fluid end”) are often truck-mounted for easy relocation from well-to-well. The fluid end comprises a housing incorporating one or more functional units, each functional unit typically comprising a suction valve, a discharge valve, and a plunger or piston bore in which a reciprocating plunger or piston alternately produces suction strokes and pressure strokes. Such pumps are being operated at unprecedented peak pumped-fluid pressures in current practice (e.g., up to about 15,000 psi), while simultaneously being weight-limited due to the carrying capacity of the trucks on which they are mounted.
Due to high peak pumped-fluid pressures, suction valves experience wide pressure variations between a suction stroke, when the valve opens, and a pressure stroke, when the valve closes. For example, during a pressure stroke a valve body may be driven toward contact with its corresponding valve seat with total valve closing force from about 50,000 to over 150,000 pounds (depending on pumped-fluid pressure and valve body area); the closing force is applied longitudinally to the proximal surfaces of the valve. Actual valve closure impact occurs with metal-to-metal contact between the valve body's valve seat interface and the valve seat itself.
Valve closure impact is particularly prominent when a conventionally-stiff valve body contacts a conventional frusto-conical valve seat. The valve body's longitudinal movement typically stops abruptly, together with the associated longitudinal movement of a proximal mass of pressurized fluid in contact with the valve body. The kinetic energy of the moving valve body and pressurized fluid is thus nearly instantly converted to a high-amplitude closing energy impulse of short duration. The effect may be compared to that of a commercially-available impulse hammer configured to produce broad-spectrum high-frequency excitation (i.e., vibration) in an object struck by the hammer.
Thus, broad-spectrum high-frequency vibration characteristically results from the closing energy impulse of a conventionally-stiff valve body contacting a conventional frusto-conical valve seat. This vibration is quickly transmitted via the valve seat to adjacent areas of the pump housing, where it can be expected to excite damaging resonances within adjacent pump housing structures. See, e.g., U.S. Pat. No. 5,979,242, incorporated by reference.
Pump housing resonances are especially problematical in the presence of corrosive components of the pumped fluid. Unsurprisingly, maintenance costs are known in the well-service industry to be relatively high and rising. Common failure modes include both rapid valve wear and the early emergence of structurally significant corrosion fatigue cracks in the pump housing. Thus, increasing attention has been directed to the prominent-but-insidious role of vibration-induced corrosion fatigue. This mechanism has also been cited in analyses of well-publicized and unanticipated structural failures of important bridges.
Proposed valve designs in the past have included relatively lighter valve bodies comprising one or a plurality of interior cavities. See, e.g., U.S. Pat. No. 7,222,837 B1, incorporated by reference. Notwithstanding the somewhat lower closing energy impulse amplitudes theoretically associated with such lighter valve bodies, they have been less popular than heavier and substantially more rigid valve bodies. The latter valve bodies have historically been relatively durable, but that performance record was largely created in lower pressure applications where the vibration issues described above are less prominent.
Problems evident during the present transition from lower pressure pump operations to higher pressures might be analogous in part to issues associated with the transition from slow-turning two-cylinder automobile engines to higher-speed and higher-powered inline six-cylinder engines. The latter transition took place around the years 1903-1910, and the new engine failure modes which became evident were neither anticipated nor understood at the time. Whereas earlier engines had been under-powered but relatively reliable, torsional crankshaft vibrations in the six-cylinder engines caused objectionable noise and vibration (“octaves of chatter from the quivering crankshaft”), as well as unexpected catastrophic failures (e.g., broken crankshafts). (Quotation cited on p. 13 of Royce and the Vibration Damper, Rolls-Royce Heritage Trust, 2003). The vibration problems, though never entirely eliminated, were finally reduced to manageable levels after several crankshaft redesigns and the development of crankshaft vibration dampers by Royce and Lanchester. Analogously, new valve designs directed to reducing the above-described adverse vibration-related effects are needed now.